4949r 2009 American Chemical Society pubs.acs.org/EF

Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452qPublished on Web 09/02/2009

Effect of Varying Compression Ratio on a Natural Gas SI Engine Performance

in the Presence of EGR

Amr Ibrahim* and Saiful Bari

Sustainable Energy Centre, School of Advanced Manufacturing and Mechanical Engineering,University of South Australia, SA 5095, Australia

Received May 14, 2009. Revised Manuscript Received August 20, 2009

The use of an exhaust gas recirculation (EGR) strategy is economically capable of satisfying theincreasingly restricted emission standards. However, the use of EGR in natural gas spark ignition (SI)engines has not been fully optimized yet. In this paper, the effects of change of compression ratio (rc) on theperformance and NO emissions of a natural gas SI engine were experimentally investigated for differentEGRdilution conditions. It was found that the use of EGRdilutionwith a stoichiometric air-fuel mixturesuppressed both surface ignition and engine knock and improved engine stability at higher compressionratios. Also, the increase of EGR dilution led to a significant reduction in NO emissions. NO emissiondecreased by about 70% when EGR dilution increased from 0 to 10% at rc = 10. In addition, engineperformance was significantly improved when the EGR dilution strategy was employed at highercompression ratios. The increase of rc from 8 to 12 at an EGR dilution of 10% increased engine brakepower by about 11% and decreased engine fuel consumption by about 10%.

1. Introduction

Recently, environmental and economical concerns havemotivated governments and research organizations to investi-gate different types of engine fuels that canbewidely used as analternative to the dominant conventional fuels of both petroland diesel. Alternative fuels are expected to be friendlier to theenvironment and more sustainable than conventional fuels.

Alternative fuels can include biofuels such as alcohols andbiodiesel, hydrogen, LPG, and natural gas. Each has itsown characteristics regarding availability, cost of production,and its effect on engine performance. Hydrogen might beconsidered as the cleanest fuel among these alternative fuels asits combustion products consist mainly of water vapor andnitrogen oxides. However, currently hydrogen is an energycarrier rather than an energy source. In other words, there aretwo main common ways for hydrogen production, waterelectrolysis using electricity and reformingofnatural gas usingsteam.Hence, the use of hydrogenas a fuelwill be limited untilan economical method of creating and distributing largequantities of it is found.

Biofuels such as alcohols and biodiesel are essentiallyderived from crops such as corn, wheat, sugar cane, soy bean,etc. Biofuels can reduce net carbon dioxide emissions as theCO2 emitted by the engine is consumed by the planted cropsused to make biofuels. However, the production of biofuelsfrom crops has become one of the main reasons that led toa global increase in food prices. Considering the limitedresources of both agricultural land andwater, the dependenceon biofuel as a sustainable fuel becomes questionable.

On the other hand, natural gas can be considered as thecleanest fossil fuel. As natural gas consists basically ofmethane, the hydrogen to carbon mole ratio (H/C) in mostnatural gas compositions is close to 3.8, which is the highest

hydrogen to carbon ratio of any hydrocarbon fuel. Thecarbon mass percentage in natural gas is close to 75%compared to 86-88% for both petrol and diesel; naturalgas produces the least CO2 per unit of energy released.1

Studies have shown that engines running on natural gas havesignificantly lower emissions than engines running on con-ventional fuels. For instance, Baldassari and co-workers2

compared natural gas and diesel engine emissions. Theauthors showed that spark-ignition (SI) natural gas engineemissions of total hydrocarbon (THC), NOx, and PM weresignificantly lower than that for a diesel-fuelled engine withemissions reductions of 67, 98, and 96%, respectively. Also,Catania and co-workers3 showed that natural gas engineemissions have a smaller impact on global warming thanpetrol emissions. Taking the global warming potential of themethane into account, the authors concluded that the naturalgas fuelled engine showed a carbon dioxide equivalent reduc-tion of 15-24% with respect to petrol.

Jang andLee4modified a petrol engine to run onnatural gasby installing a gas injection system. The authors compared theemissions from both the petrol and natural gas engines at thesame speedand load.The results showed that theuseof naturalgas fuel resulted in a reduction in HC, CO, NOx, and CO2

emissions by 49, 8, 28, and 22%, respectively, compared topetrol. Also, Gaffney and Marley5 demonstrated that a studymade by Gabele6 which compared vehicles using compressed

*To whom correspondence should be addressed. Telephone: þ6188302 5123. Fax: þ618 8302 3380. E-mail: amralihi@yahoo.com.

(1) Maclean, H. L.; Lave, L. B. Prog. Energy Combust. Sci. 2003, 29,1–69.

(2) Baldassarri, L. T.; Battistelli, C. L.; Conti, L.; Crebelli, R.; DeBerardis, B.; Iamiceli, A. L.; Gambino, M.; Iannaccone, S. Sci. TotalEnviron. 2006, 355, 64–77.

(3) Catania, A. E.; Ambrosio, S.; Mittica A.; Spessa, E. ExperimentalInvestigation of Fuel Consumption and Exhaust Emissions of a 16 VPent-Roof Engine Fueled by Gasoline and CNG, SAE, 2001, paper no.2001-01-1191.

(4) Jang, C.; Lee, J. Automobile Eng. 2005, 219, 825–831.(5) Gaffney, J. S.; Marley, N. A. Atmos. Environ. 2009, 43, 23–36.(6) Gabele, P. A. Air Waste Manage. Assoc. 1995, 45, 770–777.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

natural gas, LPG, methanol, ethanol, and petrol fuels showedthat both the LPG and compressed natural gas vehicles havethe lowest emissions of CO, benzene, aldehydes, and totalreactive volatile organic compounds.

Although natural gas is not renewable, natural gas is moresustainable than petroleum oil since the world reserves ofnatural gas are bigger than oil. Some natural gas reservesexist in parts of the world with poor oil reserves, includingAustralia. Using natural gas as an affordable energy source inthese countries can reduce their dependence on imported oil.

The lean burn technique, which utilizes excess air duringcombustion, has been widely used as a combustion techniquefor SI engines. Lean burn results in lower in-cylinder tem-perature and consequently lower thermal stresses and lowerknocking tendency. This permits the use of turbocharging,higher compression ratio (rc), and optimum spark advancetiming to achieve high engine efficiency with acceptabledurability. This technique satisfied the previous governmentalemission standards without the need for exhaust gas after-treatment, resulting in lower engine costs. As a result, most ofthe research and investment were directed toward the deve-lopment of lean burn SI engines.

Increasingly stringent ambient air quality standards areforcing the production of engines with lower emissions; seeTable 1.7 In order for the engine under the lean burn mode toproduce lower NOx emissions, it has to operate with a leanermixture. However, the use of excessive air dilution candeteriorate engine stability, increase hydrocarbon (HC) andCO emissions, and decrease engine efficiency. Another way tocontrol NOx emissions is to retard the spark timing, whichalso leads to a decrease in engine efficiency and an increase inHC emissions. Consequently, it would be difficult for theconventional gas engine operating on lean burnmode tomeetthe stringent future emission standards, especially for NOx

emissions without using exhaust gas after-treatment.The current technologies used for NOx emission after-

treatment in lean burn engines such as the selective catalyticreduction (SCR) devices are expensive and add some com-plexity to engine use. The SCR technique consists of ammoniastorage, feed, and injection system in addition to a catalyst.The ammonia is injected in the exhaust gases upstream of thecatalyst. In order for this system to operate properly, a fixedexhaust gas temperature range must be maintained.8 In addi-tion, an oxidation catalyst is necessary to reduce both the HCand CO emissions.

In order for the engines to meet future emission standards,some alternative techniques must be investigated and deve-loped.One of these techniques is the use of a three-way catalyst(TWC) to reduce NOx, HC, and CO emissions. The TWC iscapable of reducing the three emissions at the same time and isless expensive than the SCRdevices used in lean burn engines.However, in order for the TWC to operate efficiently, theengine must operate with almost a stoichiometric air-fuelmixture (i.e., without excess air). When the engine operates

near the stoichiometric mixture, the in-cylinder temperatureincreases, and consequently, the thermal stresses and theknocking tendency increase. To reduce the in-cylinder tem-perature, an inlet charge dilution must be employed. One ofthemethods used todilute the inlet charge is to recycle someofthe exhaust gases back into the cylinder intake with the inletmixture. This method is called exhaust gas recirculation(EGR). Adding EGR to the inlet mixture reduces the oxygenpartial pressure in the inlet mixture, and consequently the in-cylinder NOx production decreases.9,10 Also, the use of EGRwith a stoichiometric air-fuel mixture can economicallyreduce engine emissions by allowing the use of a TWC. Huand co-workers11,12 experimentally investigated the effect ofusing EGRdilution on the performance and emissions of a SIengine fuelled by natural gas-hydrogen blends. The authorsused a TWC during their research and they measured theengine NOx, HC, and CO emissions after the catalyst. Theauthors demonstrated that the use of the TWC with EGRdilution has the capability of satisfying the Euro4 emissionstandard.11

Several investigations that compared the effects of the useof lean combustion and EGR strategies on natural gas enginefuel consumption indicated that the stoichiometric air-fuelmixture with EGR strategy resulted in higher fuel consump-tion compared to the lean burn strategy.13-15 In 2007,Saanum and co-workers15 concluded that a penalty in enginethermal efficiency must be accepted when EGR is used as analternative to lean burn. However, these results indicate thatthe use of a stoichiometric air-fuel mixture with EGR innatural gas spark-ignition engines has not yet been fullyoptimized. Further research is still needed to optimise naturalgas engine operation for the EGR strategy in order to achievelow emissions accompanied by high engine power and effi-ciency. The aim of this research is to experimentally investi-gate the effect of change of compression ratio on theperformance and NO emissions of a natural gas SI engineemploying the EGR technique.

Table 1. The European Emission Standards, g/kWh7

year standard CO HC NOx PM

1996 Euro2 4 1.1 7 0.152000 Euro3 2.1 0.66 5 0.12005 Euro4 1.5 0.46 3.5 0.022008 Euro5 1.5 0.46 2 0.02

Table 2. Ricardo Engine Specifications

item value

serial No. 120/73No. of cylinders 1bore (mm) 76.2stroke (mm) 111.125capacity (cc) 507maximum speed (rpm) 3000inlet valve opens (� BTDC) 9inlet valve closes (� ABDC) 34exhaust valve opens (� BBDC) 43exhaust valve closes (� ATDC) 8

(7) Rabl, A. J. Transport. Res., D 2002, 7, 391–405.(8) Control Technologies for Hazardous Air Pollutants; EPA: 1991;

Handbook 625/6-91-014.

(9) Ibrahim, A.; Bari, S. Fuel 2008, 87, 1824–1834.(10) Ibrahim, A.; Bari, S.; Bruno, F. A Study on EGR Utilization in

Natural Gas SI Engines Using a Two-Zone Combustion Model; SAE,2007; paper No. 2007-01-2041.

(11) Hu, E.; Huang, Z.; Liu, B.; Zheng, J.; Gu, X.; Huang, B. Int. J.Hydrogen Energy 2009, 34, 528–539.

(12) Hu, E.; Huang, Z.; Liu, B.; Zheng, J.; Gu, X. Int. J. HydrogenEnergy 2009, 34, 1035–1044.

(13) Einewall, P.; Tunestal, P.; Johansson, B. Lean Burn Natural GasOperation vs. Stoichiometric Operation with EGR and a Three WayCatalyst; SAE, 2005; paper No. 2005-01-0250.

(14) Reppert, T.; Chiu, J. Heavy Duty Waste Hauler with ChemicallyCorrect Natural Gas Engine Diluted with EGR and Using a Three-WayCatalyst; National Renewable Energy Laboratory: USA, 2005; report No.NREL/SR-540-38222.

(15) Saanum, I.; Bysveen, M.; Tunestal, P.; Johansson, B. Lean BurnVersus Stoichiometric Operation with EGR and 3-Way Catalyst of anEngine Fueled with Natural Gas and Hydrogen Enriched Natural Gas;SAE: 2007; paper No. 2007-01-0015.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

2. Experimental Setup

The experimental research was carried out using a singlecylinder SI variable compression ratio Ricardo engine whosespecifications are shown in Table 2. The Ricardo engine wascoupled to an electrical dynamometer. The engine operated withpetrol fuel before it was converted to natural gas.

All the experimental research was conducted at the wide openthrottle condition (full load) while the engine speed was kept at1500 rpm. The engine load was controlled using an electricaldynamometer. The load was applied to the engine for a certainEGR dilution condition so that the engine speed was kept at1500 rpm. The change in EGR dilution resulted in a change inengine speed, therefore, the engine loadwas readjusted so that theengine speed remained constant at 1500 rpm.

Figure 1 shows a schematic of the experimental setup. Therecycled exhaust gaswas taken fromahole located on the exhaustpipe with the help of a small suction pump. The hot exhaust gaswas cooled by passing it through a water-cooled heat exchanger.Both a regulating valve and an orifice flow meter were installeddownstream from the heat exchanger in order to regulate andmeasure the exhaust gas flow, respectively. Natural gas wassupplied to the engine intake from a pipe line supply at atmo-spheric pressure.Air, natural gas, and cooled recycled exhaust gaswere mixed before they were introduced into the engine cylinder.

K-type thermocouples were used to measure the EGR coolerinlet and outlet temperatures with an accuracy of (1 oC. Enginespeed was measured with an accuracy of ( 10 rpm using amechanical tachometer. An Alcock model 450 V viscous flowair meter was used to measure the air flow rate flowing to theengine with an accuracy of (0.09 m3/h while a Dwyer modelRMCgas flowmeterwas used tomeasure the natural gas flow rateflowing into the engine with an accuracy of(0.045m3/h. Also, a 6mm square orifice meter was used to measure the flow rate of theexhaust gases recycled back to the engine intake. The pressuredifference across the orifice meter was measured with an accuracyof (0.01 kPa using a U-tube manometer. The percentage of

exhaust gases recycled back to the engine intake was calculatedas a percentage of the total inlet mass flow rate:

%EGR ¼:mEGR

:maþ :

mfþ :mEGR

� 100 ð1Þ

where _ma, _mf, and _mEGR are the flow rates of air, fuel, and EGR,respectively.

NO emissions weremeasured using Beckman Industrial model951A NO/NOx emission analyzer. This analyzer uses the chemi-luminescence technique, which depends on the emission of light,for measuring NO/NOx emissions. NO emissions were measuredwithout using a catalyst.

The setup for collecting the in-cylinder pressure crank angledata is also shown in Figure 1. The in-cylinder pressure crank-angle data were determined using water-cooled piezoelectricKistler 7061B pressure sensor, TDC position optical sensor,and magnetic pickup shaft encoder. A Kistler 5007 chargeamplifier was used to convert the output electrical charge fromthe piezoelectric pressure sensor into DC voltage. The outputsignals from the charge amplifier, TDCposition sensor, and shaftencoderwere received by an analogue digital converter in order tochange the continuous signals into digital signals that werehandled by a laptop via a data acquisition card as indicated inFigure 1. Lab View 7.1 computer software was used to monitorand save the outflow data from the charge amplifier, shaftencoder, and TDC position sensor.

The maximum error in air, fuel, and EGR flow rate measure-ments was found to be 0.7, 3, and 3.3%, respectively. Themaximum error in engine brake power was calculated from eq 2and was found to be 2.2%.

%ΔBP

BP¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiΔT

T

� �2

þ ΔN

N

� �2s

� 100 ð2Þ

where BP is engine brake power,ΔBP is the error in brake powercalculation,T is engine torque,N is engine speed, andΔT andΔN

Figure 1. A schematic of the experimental setup.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

are the errors in engine torque and speed, respectively, (ΔT =(0.4 N m and ΔN= (10 rpm).

The maximum error in engine brake specific fuel consumptionwas calculated from eq 3 and was found to be 3.7%.

%Δbsfc

bsfc¼

ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiΔ

:mf:mf

� �2

þ ΔBP

BP

� �2s

� 100 ð3Þ

where bsfc is engine brake specific consumption, andΔbsfc is theerror in engine brake specific fuel consumption calculation.

3. Results and Discussion

In this section, diluting the inlet mixture with EGR atvarious percentages of dilution is investigated at rc values of8, 10, and 12 in order to show the influence of the change of rcon engine performance and NO emissions in the presence ofEGR.Table 3 shows the operating conditions used during thisinvestigation.

3.1. Surface Ignition and Engine Knock. Employing astoichiometric air-fuel mixture with no EGR dilution at rc= 8 did not result in any abnormal combustion occurrence.However, abnormal combustion was identified when thestoichiometric air-fuel mixture with no EGR dilution wasemployed at rc of both 10 and 12. High cycle-to-cycle varia-tions occurred at this condition and the engine was difficult tostabilize at the desired load and speed when a stoichiometricair-fuel mixture with no EGRdilution was employed at rc=12. Figure 2 shows a number of pressure cycles for a stoichio-metric air-fuelmixturewithnoEGRdilution used at rc=12.Figure 2 identifies both preignition and engine knock occur-rence that took place for this condition.

Figure 3 shows a preignition pressure cycle that took placeat rc = 12 without EGR dilution. The figure shows that themaximum pressure occurred at almost TDC, which indicated

that the combustion started too early. Preignition was identi-fied when the combustion analysis was performed for thepressure cycle shown in Figure 3 and it was found that thecombustion started earlier than the spark timing, see Figure 4.Figure 4 shows that the heat release rate (HRR) started torapidly increase at about 90� BTDC while the spark timingwas set at 47� BTDC. When the engine was operated with astoichiometric air-fuel mixturewithout EGRdilution at rc of10 and 12, the combustion pressure and temperature exces-sively increased causing some overheated spots inside thecylinder which ignited the air-fuel stoichiometric mixtureearlier during some of the combustion cycles. This phenom-enon is called preignition or surface ignition. Surface ignitionis the ignition of the air-fuelmixture by anymeans other thanthe spark discharge such as an overheated valve or spark plug.Following surface ignition, a turbulent flame develops at eachsurface ignition location and propagates across the combus-tion chamber in an analogous manner to what occurs withnormal spark ignition.16

Figure 4 shows that there is a noticeable large portion ofnegative HRR near the end of combustion. This indicatesthat a large amount of cylinder heat transfer occurred due tothe excessive in-cylinder temperature caused by preignition.

Table 3. Engine Operating Conditions

item value

speed, rpm 1500spark timing, � BTDC 47cooled EGR temperature, �C 65inlet pressure, kPa 101inlet temperature, �C 27air-fuel ratio stoichiometric

Figure 2. In-cylinder pressure data recorded for a stoichiometricair-fuel mixture with no EGR dilution at rc = 12.

Figure 3. A preignition pressure cycle.

Figure 4.HRRcalculated for an abnormal combustion cycle at rc=12 without EGR.

(16) Heywood, J. B. Internal Combustion Engine Fundamentals,McGraw-Hill: USA, 1988; ISBN: 0-07-028637-X.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

The net heat release rate is the difference between the grossheat release rate resulting from burning the fuel and thecylinder heat transfer. When the cylinder heat transferbecame higher than the gross heat release rate near the endof combustion, the net heat release became negative asshown in Figure 4.

On the other hand, engine knock was identified for somecycles by the occurrence of pressure fluctuations near themaximum pressure condition as shown in Figure 5. Thesmall amplitude of the pressure fluctuations indicates theoccurrence of slight knock. Engine knock occurs when aportion of the air-fuel mixture ahead of the propagatingflame (the end gas) is autoignited due to high in-cylindertemperature and pressure. When knock occurs, much of thechemical energy of the end gas is released rapidly causinghigh local pressures and the propagation of pressurewaves ofsubstantial amplitude across the combustion chamber. Bothsurface ignition and knock can lead to engine instability andmay cause engine damage.

Retarding the spark timing from47 to 35�BTDCwhen thestoichiometric air-fuel mixture with no EGR dilution wasemployed did not prevent abnormal combustion. However,abnormal combustion started to disappear when the inletmixture was diluted with EGR at rc values of 10 and 12. Inaddition, engine stability improved with EGR dilution.Abnormal combustion disappeared when the inlet mixturewas diluted with 5 and 8% EGR dilution at rc values of 10and 12, respectively. This indicates that more EGR dilutionis needed at higher engine compression ratios in order toreduce the higher in-cylinder thermal and mechanical stres-ses and prevent abnormal combustion. The presence of highspecific heat gases such as water vapor in the EGR inaddition to the increase of the total inlet mass with theincrease of the EGR dilution help reduce the combustiontemperature. Also, the increase of EGR dilution in the inletmixture decreases oxygen concentration and slows the com-bustion rate, leading to a decrease in bothmaximum cylinderpressure and temperature.

3.2. In-cylinder Pressure. Both in-cylinder mechanical andthermal stresses increase with increase of rc. Figure 6 showsthe increase of the maximum cylinder pressure with theincrease of rc with a spark timing of 47� BTDC and 8%

EGR dilution. The maximum cylinder pressure increasedfrom about 38 to 62 bar when the rc increased from 8 to 12.However, the dilution of the inlet mixture with EGR iscapable of reducing both mechanical and thermal stresses,even at high compression ratios. This can be seen by studyingthe effect of the increase of EGR dilution in the inlet mixtureon maximum cylinder pressure at different compressionratios. As shown in Figure 7, the maximum cylinder pressuredecreased from about 68 to 45 bar when the EGR dilutionincreased from about 6 to 15% at rc = 12.

3.3. Engine Stability. Figure 8 shows the influence ofvarying the EGR percentage dilution in the inlet mixtureon the coefficient of variation in indicated mean effectivepressure (COV) for different compression ratios. The COVwas calculated as follows:16

COV ¼ σimep

imeph� 100 ð4Þ

where imep is the average indicated mean effective pressurecalculated for a number of cycles, n; and σimep is the standarddeviation of the indicated mean effective pressure. Bothparameters can be calculated as follows:17

imeph ¼Xi¼n

i¼1

imepðiÞ=n ð5Þ

σimep ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi1

n-1

Xi¼n

i¼1

ðimepðiÞ-imephÞ2vuut ð6Þ

The COV defines the cycle-to-cycle variations in indi-cated work per cycle, and it can be considered as a goodindicator for engine stability determination. Heywood16

demonstrated that vehicle drivability problems usuallyresult when COV exceeds about 10%. However, otherstudies showed that engine stability starts to deterioratewhen COV rises above 5%.18

Figure 5. A trace of slight knock occurrence as identified from thein-cylinder pressure data recorded for a stoichiometric air-fuelmixture with no EGR dilution at rc = 12.

Figure 6.The effect of changing rc onmaximum cylinder pressure ata spark timing of 47� BTDC and EGR dilution of 8%.

(17) Czarnigowski, J. A Simple Method of Analysis and Modeling ofCycle-to-Cycle Variation of Engine Work - the Example of IndicatedPressure; SAE: 2007; paper No. 2007-01-2079.

(18) Allenby, S.; Chang, W. C.; Megaritis, A.; Wyszynski, M. L.Automobile Engineering 2001, 215, 405–418.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

When a stoichiometric air-fuel mixture with no EGRdilution was employed at rc = 12, the engine stabilitydeteriorated and both the desired load and speed weredifficult to maintain. The dilution of the inlet mixture with6% EGR dilution improved engine stability but could notentirely prevent abnormal combustion occurrence. This in-creased the cycle to cycle variations and led to a COV ofabout 9%.

However, when EGR dilution was increased above 6%,the COV started to decrease rapidly as the abnormal com-bustion started to disappear. The COVwas kept at low levelswith the increase of EGR dilution until the misfire limit wasreached. Misfire occurrence increased cycle-to-cycle varia-tions and consequently the COV increased rapidly at allcompression ratios.

From Figure 8, the EGR misfire limit increases with theincrease of rc. The engine tolerance to EGR dilution in-creased from about 10 to 15% when the rc value increasedfrom 8 to 12. This might be explained by the effect of theincrease of rc on residual gas fraction. Increase of the rcdecreases the residual gas fraction, which makes the enginemore tolerant of EGR dilution.

3.4. Combustion Rate and Duration. Figure 9 shows theeffect of EGR dilution on the HRR at rc = 8. The maximumHRR decreases with the increase of percentage of EGRdilution. The crank angle at which the maximum HRRoccurs is shifted away from TDC. The increase of EGRdilution in the inlet mixture decreases the in-cylinder oxygenconcentration, and consequently, it reduces the HRR. Simi-lar results were also found by Hu and co-workers.12

Figure 9 shows that the spark timing was fixed at 47�BTDC for all conditions. When the stoichiometric air fuelmixture was used without EGR, the heat release rate in-creased rapidly after the flame development period until itreached its maximum value at about 360� BTDC. On theother hand, when the EGR dilution increased from 0 to 5%,the combustion rate decreased significantly, and conse-quently, the maximum heat release rate occurred at about367�BTDC.When an EGRdilution close to themisfire limit(10%) was used, the heat release rate decreased further andthe maximum heat release rate occurred at about 380�BTDC.

The HRR in Figure 9 was integrated in order to calculatethe net heat release as shown in Figure 10. The heat-releasecrank-angle relationship has the characteristic S-shape. Theheat release starts at almost zero at the time of spark whichwas fixed at 313� (or 47� BTDC), then it increases 10-20�after the spark timing. This angle is usually called flamedevelopment angle or sometimes the ignition delay. Follow-ing the flame development duration, the heat release notice-ably increases with crank angle until it reaches its maxi-mum value where essentially almost all of the fuel chemicalenergy has been released at basically the end of combustion.Figure 10 indicates that the maximum heat release, whichessentially identifies the end of combustion, occurs later asthe percentage of EGR dilution increases.

Figure 11 shows the influence of change of rc on the totalcombustion duration at various EGR dilution conditions.The total combustion duration was calculated as the crankangle interval from the spark timing to essentially the end ofcombustion where the heat release reaches its maximumvalue. The minimum percentage of EGR dilution in the inletmixture required for normal combustion operation at rcvalues of 10 and 12 was 5 and 8%, respectively, as describedin Section 3.1. Figure 11 shows a moderate decrease in the

Figure 7. The variations of maximum cylinder pressure with EGRdilution at a spark timing of 47� BTDC and different rc.

Figure 8. The variations of COV with EGR dilution at different rc.

Figure 9. The effect of varying EGR dilution in the inlet mixture onthe net HRR at rc = 8 and spark timing of 47� BTDC.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

total combustion duration with increase in rc. This might beexplained by the decrease of the residual gas fraction inaddition to the increase of the unburnedmixture densitywiththe increase of rc, which increased the combustion rate.

3.5. Engine Power.Figures 12 and 13 show the influence ofchange of rc on engine net indicated and brake powers,respectively, at different EGR dilutions. Although the in-crease of rc from 8 to 10 increased engine power, the increaseof rc from 10 to 12 did not significantly affect engine power.Such a trend could be explained by the effect of enginecompression ratio on both compression and expansionstroke powers. The increase in rc increases both compressionand expansion stroke powers. Consequently, if the increasein the expansion stroke power is more significant than theincrease in the compression stroke power, the engine netpower increases and vice versa. The net engine powerincreases with engine compression ratio until the increasein compression stroke power becomes as significant as theincrease in the expansion stroke power, which leads to aninsignificant change in engine net power at higher compres-sion ratios.

3.6. Engine Fuel Consumption. Figure 14 shows the influ-ence of change of rc on engine brake specific fuel consump-tion at different EGR dilutions. It was described earlier thatthe use of an undiluted stoichiometric air-fuel mixture at rcof 10 and 12 increased the in-cylinder thermal stresses and ledto abnormal combustion, such as surface ignition andknock,in addition to engine instability. However, it was found thatinlet mixture dilution with EGR with minimum percentagesof 5 and 8% at rc values of 10 and 12, respectively, wasnecessary to reduce the in-cylinder thermal and mechanicalstresses and prevent abnormal combustion occurrence andimprove engine stability.

The decrease of in-cylinder temperature with the increaseof EGRdilution up to about 8 and 10%at rc values of 10 and12, respectively, reduced the cylinder heat transfer and led toa slight decrease in engine fuel consumption by about 3%compared to the use of EGRdilution of 5 and 8%at rc valuesof 10 and 12, respectively. However, the increase of EGRdilution from 8 and 10% to close to the misfire limit at rcvalues of 10 and 12, respectively, did not significantly affectengine fuel consumption as it remained relatively constant.Engine fuel consumption started to increase rapidly as the

Figure 11. The variations of the total combustion duration withEGR dilution at different rc.

Figure 12. Indicated power variations with EGR dilution at anengine speed of 1500 rpm and different rc.

Figure 13. The change of engine brake power with EGR dilution atan engine speed of 1500 rpm and different rc.

Figure 10. The effect of varying the percentage of EGR dilution inthe inlet mixture on the net heat release at rc= 8 and spark timing of47� BTDC.

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Energy Fuels 2009, 23, 4949–4956 : DOI:10.1021/ef900452q Ibrahim and Bari

EGR dilution misfire limit was reached at all compressionratios.

It can be seen from Figures 13 and 14 that the rc has asignificant influence on engine performance. For instance,the use of EGR dilution of 10% at rc = 12 instead of rc = 8increased the engine brake power by about 11% as itincreased from about 3.06 to 3.4 kW as shown in Figure 13.In addition, engine fuel consumption was reduced by about10% as it decreased from about 340 to 306 g/kWh as shownin Figure 14. This is mainly due to the increase of enginetolerance to EGR dilution with the increase of rc. Also, theincrease of rc decreases the average exhaust gas temperature,which reduces the exhaust gas thermal losses and conse-quently improves the engine thermal efficiency.

Figure 14 also shows that the percentage of change ofengine fuel consumption varies with the change of rc. Forinstance, increasing the rc from 8 to 10 with an EGR dilutionof 10% decreased the engine fuel consumption by about7.3%. On the other hand, increasing the rc from 10 to 12 atthe same percentage of EGR dilution decreased the enginefuel consumption by only 2.8%.

3.7. NO Emissions. Figure 15 shows the effect of theincrease of EGR dilution on NO emissions at differentcompression ratios. Figure 15 shows that the increase ofEGR dilution can decrease NO emissions at rc = 8, 10, and12. The increase of EGR dilution in the inlet mixture reducesboth the in-cylinder burned gas temperature and oxygenconcentration and consequently leads to a substantial de-crease in NO emissions. The use of an undiluted stoichio-metric air-fuel mixture at rc = 10 resulted in a substantiallyhigh in-cylinder temperature and consequently high NOemissions. However, diluting the inlet mixture with anEGR dilution of 10% decreased NO emissions by about70%. In addition, the use of EGR dilution with a stoichio-metric air-fuel mixture would allow the use of a TWC for

further emission reduction. Similar results were also foundby Hu and co-workers.11

Although the use of EGR dilution decreases NO emis-sions, the increase of EGR dilution increases HC emissionsdue to the decrease of burn rate.11,16 Heywood16 demon-strated that the increase in HC emission was modest forlower EGRdilution conditions.However, the increase inHCemissions became more significant at higher EGR dilutionconditions as partial burning and misfire started to occur.

4. Conclusions

The effects of change of rc on the performance and NOemissions of a natural gas SI engine were experimentallyinvestigated for different EGR dilution conditions. Thefollowing conclusions were obtained:

(1) EGR dilution prevented surface ignition and engineknock at rc = 12.

(2) The increase of EGRdilution in the inletmixture leadsto a significant reduction in NO emissions. For in-stance, the increase of EGR dilution from 0 to 10% atrc = 10 decreased NO emissions by about 70%.

(3) The engine tolerance to EGR increases with theincrease of rc. The maximum tolerable EGR dilutionlimit increased from about 10 to 15% when rc in-creased from 8 to 12.

(4) The use of an undiluted stoichiometric air-fuel mix-ture at rc = 10 and 12 deteriorated engine stability.Engine stability was improved when the inlet mixturewas sufficiently diluted with EGR.

(5) Engine performance was significantly improved whenthe EGR dilution strategy was employed at rc = 10and 12. The increase of rc from 8 to 12 at an EGRdilution of 10% increased engine brake power byabout 11% and decreased engine fuel consumptionby about 10%.

Figure 15. The effect of the increase of EGR dilution on NOemissions at a spark timing of 47�BTDC, 1500 rpm, and different rc.

Figure 14. Brake specific fuel consumption variations with EGRdilution at a spark timing of 47� BTDC, engine speed of 1500 rpm,and different rc.